Reciprocating piston compressor having improved noise attenuation

ABSTRACT

A reciprocating piston compressor having a suction muffler and a pair of discharge mufflers to attenuate noise created by the primary pumping frequency in the primary pumping pulse. The suction muffler is disposed along a suction tube extending between the motor cap and the cylinder head of the compressor. The discharge mufflers are positioned in series within the compressor to receive discharge gases from the compression mechanism and are spaced one quarter of a wavelength from each other so as to sequentially diminish the problematic or noisy frequencies created during compressor operation. The motor/compressor assembly including the motor and compression mechanism is mounted to the interior surface of the compressor housing by spring mounts. These mounted are secured to the housing to define the position of the nodes and anti-nodes of the frequency created in the housing to reduce noise produced by natural frequencies during compressor operation.

BACKGROUND OF THE INVENTION

The present invention relates to reciprocating piston fluid compressiondevices such as hermetic refrigerant compressors, particularly withregard to quieting same.

Fluid compression devices such as, for example, refrigerant compressors,receive a gas at a suction pressure and compress it to a relativelyhigher, discharge pressure. Depending on the type of compression device,the work exerted on the gas in compressing it is characterized by aseries of intermittently exerted forces on the gas, the magnitude ofthese forces normally varying from zero to some maximum value. Forexample, in a cylinder of a reciprocating piston type compressor, thisforce ranges from zero at the piston's bottom dead center (BDC)position, to a maximum at or near the piston's top dead center (TDC)position, at which the pressure of the compressed gas is respectively ata minimum pressure (i.e., substantially suction pressure) and a maximumpressure (i.e., substantially discharge pressure). Some quantity of thegas is discharged from the cylinder as the piston assumes new positionsas it advances from BDC to TDC, and thus the compressed gas flowing fromthe cylinder is not at a uniform pressure. Rather, the gas which flowsfrom the cylinder, which is generally referred to as being at dischargepressure, actually has many different pressures.

Pulses of higher discharge pressure result in the compressed gas flowingfrom the cylinder, these pulses being in the portion of the flowing gaswhich leaves the cylinder as the piston approaches or reaches TDC. Asthe piston cycles in its cylinder, regular, equally distributed patternsof these pulses are created in the compressed gas which flows through aconduit, tube or line leading from the compression mechanism. Thepulsating flow of compressed gas through this discharge line may berepresented by sine waves of various frequencies and having amplitudeswhich may vary with changes in the quality of the refrigerant; thesechanges are effected by changes in refrigerant type, temperature orpressure. Pulsations at certain frequencies may be more noticeable, andthus more objectionable, than others.

Further, the nominal discharge pressure, i.e., the pressure at which thecompressed gas is generally considered to be, will also vary withrefrigerant quality. The frequency of these high pressure pulses in thecompressed gas flowing through the discharge line, however, has asubstantially constant frequency which directly correlates to the speedat which the gas is compressed in the cylinder, and the number ofcylinders in operation. This frequency is referred to as the primarypumping frequency, and is generally the lowest frequency exhibited bythe pressure pulsations in the compressed gas.

The amplitude of the pressure pulses at the primary pumping frequencytend to be the largest in the compressed gas flow. Because the primarypumping pulses are at low frequencies and large amplitudes, they areoften the primary cause of objectionable noise or vibrationcharacteristics in compressors or the refrigeration systems into whichthese compressors are incorporated. These systems normally also includeat least two heat exchangers, a refrigerant expansion device, andassociated refrigerant lines which link these components into a closedloop relationship. Pressure pulsations at other, higher frequencies haveamplitudes which are relatively smaller, but certain of these pressurepulsations may also be objectionable. Further, some objectionablepressure pulsations may establish themselves in the conduits or lineswhich convey refrigerant substantially at suction pressure to thecompression mechanism.

Substantial effort has been expended in attempting to quiet thesepressure pulses in addressing noise or vibration concerns, and it isknown to provide mufflers in the discharge or suction lines to helpresolve these issues. These mufflers may be of the expansion chambertype, in which a first refrigerant line portion opens directly into achamber, wherein the amplitude and/or frequency of at least one of thepulses may be altered, and from which the refrigerant exits through asecond line portion. Further, it is known that the discharge chamber inthe head of a reciprocating piston compressor can also serve as a typeof expansion chamber muffler. An expansion chamber type muffler of anytype is not entirely satisfactory, however, for it may cause asubstantial pressure drop in the gas as it flows therethrough, resultingin compressor inefficiency. Further, such mufflers may not providesufficient attenuation required by the application.

An alternative to an expansion chamber type of muffler is what is wellknown in the art as a Helmholtz resonator type of muffler wherein thewall of a portion of the discharge pressure line may be provided with aplurality of holes, that portion of the discharge line is sealablyconnected to a shell which defines a resonance chamber, the holes in thedischarge line providing fluid communication between the interior of thedischarge line and the resonance chamber. The size and/or quantityand/or axial spacing of these holes, and the volume of the resonancechamber, are variably sized to tune a Helmholtz resonator to aparticular frequency, and the amplitude of pulses at that frequency arethereby attenuated. Compared to an expansion chamber type of muffler, aHelmholtz muffler provides the advantage of not causing so significant apressure drop in the fluid flowing therethrough; thus compressorefficiency is not compromised to the same degree.

Although a Helmholtz resonator may be effective for attenuating theamplitude of fluid pulses having shorter wavelengths, in which case theresonator extends axially over at least a substantial portion of thepulse wavelength, prior Helmholtz resonator arrangements may not beeffective for attenuating the amplitude of fluid pulses having longerwave lengths. As mentioned above, the primary pumping frequency tends tobe rather low, the primary pumping pulses cyclically distributed over arather long wavelength. By way of the example of a single-speed hermeticreciprocating piston type compressor, the motor thereof rotates at aspeed which is directly correlated to the frequency of the alternatingcurrent (AC) electrical power which drives it. In the United States, ACpower is provided at 60 cycles/second. The electrical current isdirected through the windings of the motor stator, andelectromagnetically imparts rotation to the rotor disposed inside thestator. The crankshaft of the compressor is rotatably fixed to the rotorand drives the reciprocating piston, which compresses the refrigerant.Thus the primary pumping frequency is at or near 60 cycles per second.The speed of sound in refrigerant gas at the discharge temperature andpressure of this example is 7200 inches per second. Thus, in accordancewith the equation:c/f=λ  (1)where speed “c” is 7200 inches per second and frequency “f” is 60 cyclesper second, for the above example wavelength “λ” of the primary pumpingpulse is 120 inches. Notably, should the compressor be of the twocylinder variety, twice as many primary pumping pulses will be issuedper revolution of the crankshaft; thus λ will then be 60 inches. It canbe readily understood by those of ordinary skill in the art that simplyproviding a single Helmholtz resonator in the discharge line may belargely ineffective for attenuating the amplitude of a pulse which hassuch a long wavelength, for the point(s) of maximum pulse amplitude,which ought to be coincident with the resonator, may be too farseparated. In order for a single Helmholtz resonator to quiet a pulsehaving such a long wavelength, the resonator would be far too long tofacilitate easy packaging within the refrigerant system, let alonewithin the hermetic compressor housing.

What is needed is a noise attenuation system for a compression devicewhich effectively addresses the noise and vibration issues associatedwith pressure pulses of relatively long wavelength, such as primarypumping pressure pulses, and which overcomes the above-mentionedlimitations of previous muffler arrangements.

Typically, reciprocating piston compressors include a cylinder blockhaving at least one cylinder bore in which is disposed a reciprocatingpiston. The piston is operatively coupled, normally through a connectingrod, to the eccentric portion of a rotating crankshaft. Rotation of thecrankshaft, which may be operatively coupled to the rotor of an electricmotor, induces reciprocation of the piston within the cylinder bore.

Covering an end of the cylinder bore, in abutting contact with thecylinder block directly or through a thin gasket member disposedtherebetween, and in facing relation to the piston face, is a valveplate provided with suction and discharge ports which are both in fluidcommunication with the cylinder bore. Each of the suction and dischargeports are provided with a check valve through which gases arerespectively drawn into and expelled from the cylinder bore by thereciprocating piston as the piston respectively retreats from oradvances toward the valve plate.

The suction and discharge check valves are normally located adjacent andabut opposite planar sides of the valve plate and may, for example, beof a reed or leaf type which elastically deform under the influence ofthe gas pressure which acts thereon as the gas enters or leaves thecylinder bore through suction and discharge ports provided in the valveplate, and which are covered by the respective valves. The cylinder headis disposed on the side of the valve plate opposite that which faces thecylinder block, and in prior art compressors the head is in abuttingcontact with the valve plate, directly or perhaps through a thin gasketmember disposed therebetween. Alternatively, the valve plate-interfacingsurface of the head may be provided with a machined groove in which aseal is disposed, the seal compressed as the head is abutted to theinterfacing valve plate surface.

The cylinder head is normally a die cast aluminum or cast iron componentwhich at least partially defines separate suction and discharge chamberstherein. Suction pressure gas is introduced into the head suctionchamber through an inlet to the head; and the suction pressure gas isdrawn by the retreating piston from the head suction chamber through thesuction port of the valve plate, past the suction check valve, and intothe cylinder bore, where the gas is compressed to substantiallydischarge pressure. The discharge check valve prevents gas in thedischarge chamber from being drawn into the cylinder bore through thedischarge port of the valve plate.

Discharge pressure gas in the cylinder bore is expelled through thedischarge port of the valve plate, past the discharge check valve, andinto the discharge chamber of the head, from which it is expelledthrough the outlet of the head. The suction check valve prevents gas inthe cylinder bore from being expelled into the suction chamber of thehead through the suction port of the valve plate. As noted above, thedischarge chamber defined by the head of a reciprocating pistoncompressor may serve as a type of expansion chamber muffler. Enlargingthe volume of this chamber by including such a spacer generally improvesthe head's ability to perform as an expansion type discharge muffler andbetter attenuate noise associated with pulses carried by the compressedgas.

Moreover, a problem experienced with some reciprocating compressors,particularly those in which the discharge gas is conveyed directly fromthe head discharge chamber through interconnected conduits to a heatexchanger, is that discharge pressure gas within the head dischargechamber does not readily exit the head, resulting in a pressure buildupin the head discharge chamber during compressor operation. Consequently,the cylinder bore may not be fully exhausted of discharge pressure gasat the end of the compression cycle because the buildup of gas withinthe head discharge chamber inhibits the accommodation therein of gasbeing exhausted thereinto from the cylinder. Because gas from theprevious compression cycle has not been fully exhausted from thecylinder bore, less suction pressure gas can be drawn into the cylinderduring the next compression cycle. Thus, the efficiency of thecompressor is compromised. Moreover, the temperature of gas on thedischarge side of the system, both within the head itself and the highside of the system, may become excessively high as more and more work isexpended on the gas already at discharge pressure.

The previously preferred solution to this problem has been to enlargethe size of the head discharge chamber, thereby allowing gas which isexhausted from the cylinder bore to be more easily compressed into, andaccommodated by, the head discharge chamber. As noted above, enlargementof this chamber usually also facilitates improvements in noise quality.One approach to enlarging the head's discharge chamber has been toretool the head. This solution carries with it attendant tooling costswhich may not be insubstantial. Further, where a common head design isshared between different compressor models, a newly designed head whichsolves the problem for one model may not meet the needs (e.g., packagingrequirements) of the other model(s), thereby requiring a plurality ofhead designs to be released and maintained in inventory.

Another approach to enlarging the head's discharge chamber is to providea spacer between the valve plate and the existing head, whicheffectively enlarges the volume of the head discharge chamber (and thesuction chamber as well). The spacer comprises a separate componentwhich may be used in one compressor but not another, the two compressormodels sharing a common head design. These spacers may be made ofplastic or metal.

Previous plastic spacers have had coefficients of thermal expansionwhich differ substantially from those of the cylinder block and/or thehead, and consequently may either shrink and thereby cause a leak acrossits sealing surfaces, or expand and be overly compressed between thevalve plate and head, thereby placing considerable additional stress onthe spacer, the head and bolts which extend through the spacer andattach the head and spacer to the cylinder block. If so stressed, thespacer may crack and consequently leak. Plastic spacers do, however,provide the benefits of being lightweight, and providing insulationagainst thermal conduction between the head and the cylinder block,thereby keeping the discharge gas somewhat cooler and thus reducing thecapacity required of the heat exchanger which condenses the highpressure gas to a high pressure liquid. Plastic spacers are also madeinexpensively by injection molding techniques.

Previous metal spacers, on the other hand, undesirably promote thermalconduction between the head and the cylinder block, weigh more, andusually are die cast and machined, resulting in a relatively moreexpensive part vis-a-vis a plastic spacer. A metal spacer, however, mayhave a coefficient of thermal expansion which avoids the above mentionedshrinkage and stress concerns attendant with plastic spacers. Further,prior plastic and metal spacers alike may require additional, separategaskets to seal the opposite open spacer ends to the valve plate andhead in order to provide a proper seal.

What is needed is an inexpensively produced head spacer for increasingthe volume of the discharge chamber of the cylinder head, which providesseals between the head spacer and the valve plate, and between the headspacer and the cylinder head, without the need for additional seals.

Further, it is known to dispose an end cap over the end of the annularmotor stator in a low-side hermetic compressor, the end cap coveringboth the stator end and the end of the motor rotor disposed inside thestator. It is also known to drawn suction pressure refrigerant gas fromwithin the end cap through a suction tube extending therefrom which isin fluid communication with the inlet to a compression mechanism drivenby the motor and disposed at the opposite end of the motor stator. Sucha configuration is shown, for example, in U.S. Pat. No. 5,129,793 (Blasset al.) and U.S. Pat. No. 5,341,654 (Hewette et al.), and exemplified bythe Model AV reciprocating compressors manufactured by the TecumsehProducts Company of Tecumseh, Mich. It is also known to provide suctionmufflers in this tube intermediate the stator end cap and thecompression mechanism, as taught by Blass et al. '793 and Hewette et al.'654.

A problem with such suction tube arrangements is that their lengths arefixed and particular to stators of a given height. A unique suction tubedesign must be provided for each different stator height in compressorassemblies which might otherwise be similar, resulting in partcomplexities and associated inventorying costs and efforts, andadditional jigs and fixtures to produce different suction tubeassemblies to accommodate these various stators. It would be desirablyto provide a single suction tube assembly, with or without a mufflerprovided therein, which extends between the stator end cap and the inletto the compression mechanism and can accommodate stators of differentheights. Further, it may also be desirable to fix the distance of themuffler from the inlet to the compression mechanism to aid in properlytuning or packaging the muffler, while still accommodating thesedifferent stators.

Further still, it is known to resiliently support the motor/compressorassembly, which includes the motor and compression mechanism, within thehermetic shell or housing on a plurality of mounts affixed to theinterior of the housing. Typically, these mounts are equally distributedabout the interior circumference of the housing or otherwise placedthereabout in a manner which is merely convenient to attachment of themounts to the motor/compressor assembly.

It is further understood by those of ordinary skill in the art that thehousing has natural resonant frequencies that may produce loud, pure,undesirable tones when the housing is vibrated at or near thosefrequencies. Typically, equally distributing the mounts about the innercircumference of the housing may, at the points of contact therebetween,establish nodes which coincide with at least one of these naturalfrequencies. Similarly, placement of the mounts merely to facilitateconvenient mounting of the motor/compressor assembly may also placethese points of contact at nodes of natural frequencies which produceloud tones. Thus, previous compressors do not beneficially place themotor/compressor mounts on the housing in a manner which addresses thenoise associated with excitation of these natural frequencies. To do sowould reduce or eliminate the housing's natural resonant frequencies,and reduce the noise produced thereby.

SUMMARY OF THE INVENTION

One aspect of such a noise attenuation system for a compression devicerelates to an improved discharge pulse reduction system which comprisesat least one muffler located in a discharge fluid line, the mufflerspaced along the discharge line at a distance from a compressordischarge chamber or another upstream muffler which is a particularfraction or multiple of the wavelength of the primary pumping frequency.Thus, the amplitude of the primary pumping frequency, which may bereduced in the above-mentioned compressor discharge chamber or upstreammuffler, is further reduced by the muffler placed at the above-mentioneddistance therefrom, at which the already reduced amplitude reaches itsnew maximum value. Thus, the amplitude of the pulse at the primarypumping frequency is twice attenuated, improving the noise and vibrationcharacteristics of the compressor and/or the refrigerant system intowhich it is incorporated. The muffler(s) may be of the Helmholtz orexpansion chamber type.

Accordingly, the present invention provides a compressor assemblyincluding a compression mechanism into which a gas is receivedsubstantially at a suction pressure and from which the gas is dischargedsubstantially at a discharge pressure, the gas discharged from thecompression mechanism carrying pressure pulses having a particularfrequency and wavelength, these pressure pulses being of variableamplitude. A first muffler is provided through which the gas dischargedfrom the compression mechanism flows, and a second muffler is providedin series communication with the first muffler and through which the gashaving flowed through the first muffler flows. The first and secondmufflers are spaced by a distance which is substantially equal to an oddmultiple of one quarter of the wavelength, the amplitude being reducedin response to the gas having flowed through the second muffler.

The present invention also provides a compressor assembly including acompressor mechanism into which a gas is received substantially at asuction pressure and from which the gas is discharged substantially at adischarge pressure, the gas discharged from the compression mechanismcarrying pressure pulses having a particular frequency and wavelength,these pressure pulses being of variable amplitude. Also provided is aconduit through which gas substantially at discharge pressure flows, andmeans for reducing the amplitude of the pressure pulses at locations atwhich the amplitudes reach their highest absolute values.

The present invention further provides a method for reducing theamplitude of pressure pulses having a particular wavelength in a fluid,including: flowing the pressure pulse-containing fluid through aconduit; attenuating the pressure pulse amplitude at a first locationalong the conduit; and further attenuating the pressure pulse amplitudeat a second location along the conduit distanced from the first locationa distance which is substantially equal to an odd multiple of onequarter of the wavelength.

A head spacer is provided for increasing the volume of a dischargechamber in the cylinder head assembly of a reciprocating pistoncompressor, in which the head spacer is disposed between a valve plateand a cylinder head, and has a plurality of substantially concentric,alternating ridges and valleys disposed around the periphery of firstand second end surfaces of the head spacer. When the cylinder head istorqued down onto the cylinder block in response to a compressive loadexerted on the cylinder head during the assembly of the cylinder headassembly, the tips of the ridges deform to form a continuous labyrinthseal between the head spacer and the cylinder head, and between the headspacer and the valve plate.

The head spacer may be made from an injection-molded plastic, and has acoefficient of thermal expansion which is substantially similar to themetal components of the cylinder head assembly, such that the headspacer may shrink and/or expand at the same rate as the cylinder blockand cylinder head. Further, the plastic from which the head spacer ismade provides insulation against thermal conduction between the valveplate and the cylinder head.

In one form thereof, a reciprocating piston compressor is provided,including cylinder block having a cylinder bore; a pistonreciprocatingly disposed in the cylinder bore; a cylinder head connectedto the cylinder block and partially defining a suction chamber intowhich gas is received and from which the gas exits into the cylinderbore substantially at a suction pressure, the cylinder head partiallydefining a discharge chamber into which gas is received from thecylinder bore and from which the gas exits substantially at a dischargepressure; a valve plate having a suction port through which the cylinderbore and the suction chamber fluidly communicate, and a discharge portthrough which the cylinder bore and the discharge chamber fluidlycommunicate; a suction check valve disposed over the suction port andpast which gas flows from the suction chamber to the cylinder bore, flowfrom the cylinder bore to the suction chamber being inhibited by thesuction check valve; a discharge check valve disposed over the dischargeport and past which gas flows from the cylinder bore to the dischargechamber, flow from the discharge chamber to the cylinder bore beinginhibited by the discharge check valve; and a spacer disposed betweenthe valve plate and the cylinder head, the spacer having generallyopposite first and second end surfaces, each of the first and secondspacer and surfaces respectively abutting an interfacing surface of thevalve plate and the cylinder head, the spacer partially defining thedischarge chamber, a substantial portion of the volume of the dischargechamber located between spacer end surfaces; wherein the first andsecond spacer end surfaces are each provided with a plurality ofsubstantially concentric ridges having tips, the ridge tips having oneof a deformed state and an undeformed state, adjacent ones of the ridgesseparated by a valley, the ridge tips being placed in the deformed statein response to a compressive load exerted on the spacer between thevalve plate and the cylinder head during assembly of the compressor, thedeformed ridge tips providing a seal between the first spacer endsurface and the valve plate, and between the second spacer end surfaceand the cylinder head.

In a further form thereof, a cylinder head spacer for a reciprocatingpiston compressor is provided, including a body portion made of aplastic material and having a substantially open interior extendingbetween first and second end surfaces; and a plurality of substantiallyconcentric, alternating ridges and valleys extending around a peripheryof each of the first and second end surfaces, the ridges having one of adeformed state and an undeformed state, the ridges being placed in thedeformed state in response to a compressive load exerted on the firstand second end surfaces, such that the ridges extend into the valleysand contact adjacent ridges to form sealing surfaces, the sealingsurfaces coplanar with the first and second end surfaces.

In another form thereof, a method of assembling a reciprocating pistoncompressor having a cylinder block with a cylinder bore opening, a valveplate, and a cylinder head, is provided, including the steps ofproviding a spacer having first and second end surfaces each providedwith a plurality of substantially concentric ridges having tips, theridge tips having one of a deformed state and an undeformed state,adjacent ones of the ridge tips separated by a valley; orienting thevalve plate, the spacer, and the cylinder head in a stack arrangementover the cylinder bore opening; and exerting a compressive load on theridge tips to deform the ridge tips to the deformed state, the deformedridge tips providing sealing contact between the first spacer endsurface and the valve plate, and between the second spacer end surfaceand the cylinder head.

In a still further form thereof, a method is provided of assembling acylinder head assembly of a reciprocating piston compressor, thecompressor having a cylinder block with a bolt hole therein, includingthe steps of providing a bolt, a suction leaf plate, a valve plate, anda cylinder head, each of which include a bolt hole therein; providing aspacer having a bolt hole, and first and second end surfaces eachprovided with a plurality of continuous, alternating ridges and valleysextending around a periphery of each of the first and second endsurfaces, the ridges including tips having one of a deformed state andan undeformed state; positioning the suction leaf plate, the valveplate, the spacer, and the cylinder head, respectively, on the cylinderblock such that the bolt holes are aligned; inserting the bolt throughthe bolt holes, and tightening the bolt to exert a compressive load onthe ridge tips and deforming the ridge tips to the deformed state, thedeformed ridge tips providing sealing contact between the first spacerend surface and the valve plate, and between the second spacer endsurface and the cylinder head.

One advantage of the present head spacer is that it is inexpensivelyproduced, and, because the head spacer comprises an individualcomponent, the head spacer may used with existing compressor designswithout retooling other components of the cylinder head assembly.

Another advantage is that the labyrinth seal produced by the deformationof the ridge tips of the head spacer obviates the need for additionalseals between the head spacer and the valve plate, and between the headspacer and the cylinder head.

A further advantage is that the plastic material of the head spacer bothprovides insulation against thermal conduction between the cylinderblock and the cylinder head, and has a coefficient of thermal expansionsubstantially similar to the other metal components of the cylinder headassembly to prevent the leakage due to the shrinkage and expansion whichis observed with existing head spacers.

Another aspect of the inventive noise attenuation system for acompression device relates a suction tube assembly which extends betweenthe stator end cap and the inlet to the compression mechanism, and maybe telescoped in the general direction of the stator's longitudinal axisto accommodate stators of different heights. Certain embodiments of thissuction tube assembly include a muffler, and this muffler may have alocation which is fixed relative to the compression mechanism.

Accordingly, the present invention provides a compressor assemblyincluding a compression mechanism having an inlet into which a gassubstantially at suction pressure is received, and an outlet from whichgas compressed by the compression mechanism is discharged substantiallyat a discharge pressure. A motor is also included which includes a rotorand a stator, the stator substantially surrounding the rotor and havingan end, the rotor operatively coupled with the compression mechanism. Anend cap is disposed over the stator end, the end cap having an interiorin which is gas substantially at suction pressure. A suction tube ofvariable length is also provided through which the compression mechanisminlet and the end cap interior are in fluid communication, the suctiontube comprising first and second tubes which are in sliding, telescopingengagement, whereby the length of the suction tube may be adjustedthrough relative axial movement of the first and second tubes.

The present invention also provides a compressor assembly including acompression mechanism having an inlet into which a gas substantially atsuction pressure is received, and an outlet from which gas compressed bythe compression mechanism is discharged substantially at a dischargepressure, and a motor having a rotor and a stator selected from aplurality of stators of differing heights. The stator substantiallysurrounds the rotor and has opposite ends distanced by the stator'sheight. The rotor is operatively coupled with the compression mechanism.An end cap is disposed over one of the stator ends and has an interiorsubstantially at suction pressure, and first and second telescopinglyengaged tubes defining a suction tube which extends axially over atleast a portion of the stator height and through which the end capinterior and the compression mechanism inlet are in fluid communication.The suction tube has a length which is varied in response to therelative axial positions of the telescopingly engaged first and secondtubes, whereby the suction tube length may be varied to accommodate adifferent stator alternatively selected from the plurality of stators.

Further, the present invention provides a compressor assembly includinga compression mechanism having an inlet into which a gas substantiallyat suction pressure is received, and an outlet from which gas compressedby the compression mechanism is discharged substantially at a dischargepressure, and a motor having a rotor and a stator selected from aplurality of stators of differing heights. The stator substantiallysurrounds the rotor and has opposite ends distanced by the stator'sheight. The rotor is operatively coupled with the compression mechanism.An end cap is disposed over the stator and has an interior in which isgas substantially at suction pressure, the end cap being distanced fromthe compression mechanism inlet an amount dependent upon the stator'sheight. A tube assembly is provided through which gas is directed fromthe end cap interior to the compression mechanism inlet, the tube havingmeans for adjusting its length, whereby the compressor assembly couldalternatively comprise a different stator selected from the plurality ofstators.

Still another aspect of the inventive noise attenuation system for acompression device relates to motor/compressor assembly mounts which areattached to the interior of the compressor housing in a manner whichreduces or eliminates natural resonant frequencies of the housing. Themounts are distributed unequally about the inner circumference of thehousing and attached thereto a positions which do not coincide withnodes of these frequencies. That is, the mounts are secured to theinside of the housing to interfere with the wave form produced by thenatural frequencies in the compressor housing so as to reduceobjectionable noise. Resonation of the housing at these naturalfrequencies is thus prevented, and the compressor quieted.

Accordingly, the present invention provides a compressor assemblyincluding a housing having at least one natural frequency having a waveform with amplitude large enough for the housing, when vibrated at thatfrequency, to produce an objectionable noise. The natural frequency waveform has a plurality of natural nodes equally distributed about thecircumference of the housing and natural anti-nodes located betweenadjacent natural nodes. A motor/compressor assembly is also providedwhich includes a compression mechanism in which gas is compressed fromsubstantially a suction pressure to substantially a discharge pressure,and a motor operably engaged with the compressor mechanism. A pluralityof mounts are unequally distributed about the circumference of thehousing, the motor/compressor assembly being supported within thehousing by the mounts. Each mount is attached to the housing at a firstpoint, the first points not coinciding with the natural nodes of thenatural frequency wave form. These first points define forced nodes onthe circumference of the housing to which the nodes of the naturalfrequency wave form are forced, and the natural frequency wave form isaltered in response to the natural nodes being forced to the forcednodes, whereby the housing is prevented from vibrating at the naturalfrequency.

BRIEF DESCRIPTION OF THE DRAWINGS

The above-mentioned and other features and advantages of this invention,and the manner of attaining them, will become more apparent and theinvention itself will be better understood by reference to the followingdescription of an embodiment of the invention taken in conjunction withthe accompanying drawings, wherein:

FIG. 1 is a first longitudinal sectional view of a first embodiment of acompressor in accordance with the present invention;

FIG. 2 is a second longitudinal sectional view of the compressor shownin FIG. 1, along line 2-2;

FIG. 3 is a sectional view of the compressor shown in FIG. 1, along line3-3;

FIG. 4 is a sectional view of the compressor shown in FIG. 1, along line4-4;

FIG. 5 is a sectional view of the compressor shown in FIG. 1, along line5-5;

FIG. 6 is a bottom view of the crankcase of the compressor shown in FIG.1;

FIG. 7A is a first side view of the suction muffler of the compressorshown in FIG. 1;

FIG. 7B is a second side view of the suction muffler shown in FIG. 7A;

FIG. 7C is a third side view of the suction muffler shown in FIG. 7A, inan alternative configuration in which the inlet tube thereof isshortened;

FIG. 8A is an enlarged plan view of the valve assembly of the compressorshown in FIG. 1;

FIG. 8B is an exploded side view of the valve assembly shown in FIG. 8A;

FIG. 9A is a first plan view of a discharge tube of the compressor shownin FIG. 1, the discharge tube including a discharge muffler;

FIG. 9B is a second plan view of the discharge tube of FIG. 9A;

FIG. 10 is a first longitudinal sectional view of a second embodiment ofa compressor according to the present invention;

FIG. 11 is a second longitudinal sectional view of the compressor shownin FIG. 10, along line 11-11;

FIG. 12 is a third longitudinal sectional view of the compressor shownin FIG. 10, along line 12-12;

FIG. 13 is a sectional view of the compressor shown in FIG. 10, alongline 13-13;

FIG. 14 is a bottom view of the compressor shown in FIG. 10;

FIG. 15 is a sectional view of the compressor shown in FIG. 10, alongline 15-15, in which the motor and compression mechanism are not shown;

FIG. 16 is a sectional view of the compressor shown in FIG. 10, alongline 16-16, in which the motor, compression mechanism, bottom housing,and discharge tube are not shown;

FIG. 17 is a bottom view of the crankcase of the compressor shown inFIG. 10;

FIG. 18A is a plan view of one embodiment of the head spacer included inthe compressor shown in FIG. 10;

FIG. 18B is a side view of the head spacer shown in FIG. 18A;

FIG. 18C is a perspective view of an alternative embodiment of the headspacer included in the compressor shown in FIG. 10;

FIG. 18D is a partial sectional view of the head spacer of FIG. 18C,showing the spacer prior to installation;

FIG. 18E is a partial sectional view of the head spacer of FIG. 18D,showing the spacer installed;

FIG. 19A is a side view of the suction muffler of the compressor shownin FIG. 10;

FIG. 19B is a longitudinal sectional view of the suction muffler shownin FIG. 19A;

FIG. 20 is a longitudinal sectional view of the first discharge mufflerof the compressor shown in FIG. 10;

FIG. 21 is a view of the discharge tube of the compressor shown in FIG.10, the discharge tube including the second discharge muffler;

FIG. 22 is a longitudinal sectional view of the second discharge mufflerof the compressor shown in FIG. 10;

FIG. 23A is a schematic view of the primary pumping pulse in thedischarge refrigerant in the compressor of FIG. 10 for various distancesbetween the first and second mufflers of that compressor;

FIG. 23B is a schematic view of the amplitude of the primary pumpingpulse in the discharge refrigerant in the compressor shown in FIG. 10,after passing through the first and second mufflers spaced a distance D;

FIG. 23C is a schematic view of the amplitude of the primary pumpingpulse in the discharge refrigerant in the compressor shown in FIG. 10,after passing through the first and second mufflers spaced a distanceD′;

FIG. 24 is a perspective view of a compressor housing showing theformation of a vibration at a natural frequency; and

FIG. 25 is a sectional view of the compressor shown in FIG. 5,schematically illustrating a natural frequency wave form and a forcedfrequency wave form in the compressor housing.

Corresponding reference characters indicate corresponding partsthroughout the several views. Although the drawings representembodiments of the present invention, the drawings are not necessarilyto scale and certain features may be exaggerated in order to betterillustrate and explain the present invention.

DETAILED DESCRIPTION OF THE INVENTION

Referring to FIGS. 1 and 2 there is shown a first embodiment of areciprocating piston compressor assembly according to the presentinvention. Reciprocating piston compressor assembly 20 is a hermeticcompressor assembly which may be part of a refrigeration orair-conditioning system (not shown). Compressor 20 is a 5-ton compressorhaving a displacement of approximately 5.6 cubic inches. Compressorassembly 20 comprises housing 22 having an interior surface to whichmounts 24 are attached (FIGS. 1-5). Mounts 24 include springs whichresiliently support motor/compressor assembly 26, to vibrationallyisolate the motor/compressor assembly from housing 22 in a manner thatwill be described hereinbelow. Motor/compressor assembly 26 comprisesmotor 28 and compression mechanism 30. In the depicted embodiment,compression mechanism 30 is of the reciprocating piston type, althoughit is to be understood that certain aspects of the present invention maybe adapted to other types of compressor assemblies. Previousreciprocating piston compressors are described in U.S. Pat. No.5,224,840 (Dreiman et al.) and U.S. Pat. No. 5,951,261 (Paczuski), thedisclosures of which are expressly incorporated herein by reference.These incorporated patents are assigned to the assignee of the presentinvention.

Motor 28 comprises stator 32 which is provided with windings 33, androtor 34 as illustrated in FIG. 2. Alternating current from an externalpower source (not shown) is directed through stator windings 33 viaterminal cluster 35 (FIGS. 3, 4 and 5) to electromagnetically inducerotation of rotor 34. Crankshaft 36 extends longitudinally throughcentral aperture 37 in rotor 34 to which it is rotatably attached todrive compression mechanism 30. Shaft 36 is operably coupled to a pairof pistons 38 which are reciprocatively disposed in cylinder bores 40formed in cylinder block 41 of cast-iron crankcase 42, which is attachedto the lower one of two opposite ends of the stator.

During compressor operation, refrigerant at suction pressure is drawninto housing 22; compressor assembly 20 is a low-side compressor, motor28 being in a low pressure and low temperature environment. The suctionpressure refrigerant is drawn into housing 22 through inlet 45 which isheld securely within aperture 47 located in the side of housing 22 bywelding, brazing or the like (FIG. 3). As illustrated in FIG. 3, inlet45 is substantially aligned with suction inlet 46 located in one side ofmotor end cap 44 such that as suction pressure refrigerant is drawn intohousing 22, a portion of the fluid enters motor end cap 44 through inlet46. The remainder of the suction pressure fluid circulates withinhousing 22. The suction pressure refrigerant which flows into motor endcap 44, flows over the top of motor 28 to cool the top end thereof. Therefrigerant exits motor end cap 44 through suction tube 48 which leadsto inlet 50 of suction muffler 52. Suction muffler 52 is a steel,expansion type muffler shown in FIGS. 7A-7C and includes expansionchamber 54 having a volume of 3.531 cubic inches. Alternatively, suctionmuffler 52 may be modified such that its expansion chamber 54 has avolume of 4.63 cubic inches. Suction muffler 52 has inlet 50 and outlet56 which are sealingly connected to suction tubes 48 and 58,respectively (FIGS. 7A-7C). Suction tubes 48 and 58 have a diameter ofapproximately ⅞ inch and along with muffler 52 are constructed from amaterial such as steel. Although the openings suction tubes 48 and 58are shown as being substantially offset within expansion chamber 54(FIG. 7A), muffler 52 may be modified to more closely align theseopenings so that fluid may flow more directly between them withinchamber 54. Moreover, those of ordinary skill in the art will recognizethat the extent to which the ends of tubes 48 and 58 extend into chamber54 may vary considerably depending on the frequency being attenuatedwithin the muffler.

As shown in FIGS. 1 and 2, suction tube 58 is received in one end ofsuction plenum 60 which is secured at end 62 to cylinder head inlets 64of cylinder head 66. Suction plenum 60 is a plastic insert into whichsteel tube 58 is interference fitted and is held in place over suctionchamber 61 in cylinder head 66 by strap 68. Suction muffler 52 is tunedto attenuate noises created by suction check valves and pressure pulseshaving a frequency between 1000 and 1400 hertz.

Referring to FIGS. 7A-7C, in the shown embodiment of the presentinvention, suction tube 48 includes first tube 70 and second tube 72 inwhich the outer and inner diameters, respectively, are telescopicallyengaged. Suction tube 48 is constructed from steel, but may beconstructed from any suitable material to withstand the compressorenvironment. First tube 70 has an outer diameter of ⅞ inch and is of aslightly smaller diameter than second tube 72, which has an outerdiameter of 1 inch. A sealing member such as O-ring 73 is disposedbetween first tube 70 and second tube 72 so as to sealingly engage theinner surface of second tube 72 with the outer surface of first tube 70.First portion 70 is then telescopically movable within second tube 72 toprovide an adjustable suction tube 48 having different lengths toaccommodate different stator heights H (FIG. 2), i.e., the distancebetween the opposite ends of a stator. The position of muffler 52 issuch that tubes 70 and 72 are axially aligned along the generaldirection of the stator height.

As shown in FIG. 2, from suction muffler 52, suction pressure gas isintroduced into suction plenum 60 and into suction chamber 61 ofcylinder head 66 from which the gas is drawn by the retreating pistons38 through the suction check valves of valve assembly 74 (FIGS. 8A and8B), and into cylinder bores 40, wherein the gas is compressed tosubstantially discharge pressure. Cylinder head 66 is a material such ascast iron or aluminum. Once compressed, the discharge pressure gasesflow past the discharge valve of valve assembly 74 and into dischargechamber 76 defined within cylinder head 66. Discharge chamber 76 of thisembodiment is of a size which is great enough to act as an expansiontype muffler wherein the amplitude of the pressure wave of thecompressed fluid is altered, thereby attenuating the noise created bythe operation of the discharge valves and the primary pumping frequency.The volume of discharge chamber 76 is 6.93 cubic inches.

In the usual fashion, valve assembly 74 is provided between crankcase 42and cylinder head 66 to direct the suction pressure and dischargepressure gases into and out of cylinder bores 40. Valve assembly 74 isillustrated in FIGS. 8A and 8B and includes valve plate 78 havingcentrally located suction ports 80 and surrounding discharge ports 81shown in dashed lines in FIG. 8A. Discharge ports 81 are disposedbeneath retaining plate 82. Valve plate 78 and retaining plate 82 areconstructed from a material such as steel. Between valve plate 78 andretaining plate 82 are discharge check valves 84 which open and closedischarge ports 81. Discharge check valves 84 are made from springsteel, as is well known in the art. Each discharge check valve 84prevents gas in discharge chamber 76 from being drawn into a cylinderbore 40 through the associated discharge ports 81 of valve plate 78.Discharge pressure gas in cylinder bores 40 is expelled through thedischarge ports of valve plate 78, past discharge check valves 84, andinto discharge chamber 76, from which it is expelled through outlet 86of cylinder head 66 into discharge tube 88 (FIGS. 1 and 3).

Positioned on the opposite side of valve plate 78 are a pair of pins 90which are aligned across suction ports 80 and fixed to valve plate 78.Thin metal suction check valves 92 are constructed from spring steel asare discharge check valves 84 and include a pair of slots 93, one beingdisposed at opposite ends of valves 92 (FIG. 8B). Suction check valves92 are positioned so that pins 90 are received within slots 93 to guidevalves 92 between open and closed position. Suction valves 92 preventgas in cylinder bores 40 from being expelled into suction chamber 61 incylinder head 66 through suction ports 80 of valve plate 78. In thisparticular compressor 20, two valve assemblies 74 are provided on commonplate 78, one valve assembly 74 being disposed over each cylinder bore40.

The discharge pressure gases in discharge 76 are directed into adischarge tube which, as shown, may be comprised of multiple,series-connected tubes. The discharge tube extends from head 66 throughaperture 94 in housing 22, and is connected to the remainder of therefrigerant system (FIGS. 1, 3, and 4). This housing aperture is sealedabout the discharge tube by any suitable manner. Referring now to FIGS.9A and 9B there is shown discharge tube 96 which comprises part of thecompressor discharge tube assembly. Discharge tube 96 is somewhatflexible in nature so that shocks associated with pressure pulses may beabsorbed by the resilient flexing of tube 96. Discharge tube 96 issecured to discharge tube 88 at 98 by any suitable method such aswelding or brazing (FIG. 1).

Located along discharge tube 96 is expansion type discharge muffler 100which is a second muffler of compressor assembly 20 for further reducingthe undesirable noise in the refrigerant gas. Flow of compressedrefrigerant gas is directed along discharge tube 88 and discharge tube96 in the direction of arrows 102 through muffler 100 (FIGS. 1, 9A and9B). Both discharge tube 88 and discharge tube 96 are approximately ½inch in diameter and are formed from a material such as steel. Muffler100 is specifically spaced from discharge chamber 76 in cylinder head 66in accordance with the present invention as will be describedhereinbelow.

Referring to FIGS. 10, 11 and 12, compressor assembly 104 is a secondembodiment of a reciprocating piston compressor assembly according tothe present invention. Compressor 104 is a 3-ton compressor having adisplacement of approximately 3.5 cubic inches. Compressor assembly 104is similar in structure and operation to compressor assembly 20 exceptas described herein. Suction pressure gases enter compressor housing 22through inlet 45 which is held securely within aperture 47 located inthe side of housing 22 by welding, brazing or the like. As illustratedin FIGS. 10 and 13, inlet 45 is substantially aligned with suction inlet46 located in one side of motor end cap 44 such that as suction pressurerefrigerant is drawn into housing 22, a portion of the fluid entersinlet 46 into motor end cap 44. The remainder of the suction pressurefluid circulates within housing 22. The suction pressure refrigerantwhich flows into motor end cap 44, flowing over the top of motor 28 tocool the top end thereof. The refrigerant exits motor end cap 44 throughsuction tube 48 which leads to inlet 50 of suction muffler 106 as shownin FIGS. 12, 19A, and 19B. Suction muffler 106 is of a Helmholtz typehaving tube 108, which is part of suction tube 48, provided with aplurality of axially-spaced hole arrangements 110 therealong. Tube 48 isconstructed from a material such as steel or the like and has a diameterof approximately ¾ inch. Each arrangement of holes 110 comprises twopairs of holes 112, the holes in each arrangement are cross drilled sothat the holes are equally radially distributed about the circumferenceof tube 108. Notably, each hole arrangement 110 is substantially equallyspaced along the longitudinal axis of tube 108. The number and size ofholes 112 is dependant on the frequencies which are being attenuated. Inthis embodiment, holes 112 are formed in tube 108 by any suitable mannersuch as being punched or drilled and have a diameter of 3/16 inch toattenuate noise created by the operation of valve arrangement 74 and theprimary pumping frequency. Tube 108 is surrounded by shell 114 havingends 116 and 118 which are sealed to the exterior surface of tube 108 tocreate chamber 120 around hole arrangements 110 (FIG. 12, 19A, and 19B).Shell 114 is made from any suitable material such as steel and has avolume of 1.16 cubic inches which is also dependant on the frequenciesin the primary pumping pulse being attenuated.

As with compressor assembly 20, the suction gas exits suction muffler106 and enters cylinder head assembly 122 which includes cylinder head66 covering a head spacer disposed between valve plate 78 and cylinderhead 66, as described in more detail below (FIG. 12).

Cylinder head 66 and the head spacer together define enlarged suctionchambers 126 and discharge chamber 128 therein which help to alleviateefficiency problems experienced with some reciprocating compressors.These problems include discharge pressure gas within discharge chamber128 not readily exiting cylinder head 66, resulting in a pressurebuildup in discharge chamber 128 during compressor operation.Consequently, cylinder bore 40 may not be fully exhausted of dischargepressure gas at the end of the compression cycle because the buildup ofgas within discharge chamber 128 inhibits the accommodation therein ofgas being exhausted thereinto from cylinder 40. Because gas from theprevious compression cycle has not been fully exhausted from cylinderbore 40, less suction pressure gas can be drawn into cylinder 40 duringthe next compression cycle. Thus, the efficiency of the compressor iscompromised. Moreover, the temperature of gas on the discharge side ofthe system may become excessively high as more and more work is expendedon the gas already at discharge pressure.

A first embodiment of head spacer 124 is shown in FIGS. 18A and 18B andprovides means for enlarging suction chamber 126 and discharge chamber128 (FIG. 12). Spacer 124 includes body portion 130 having asubstantially open interior extending between first planar end surface132 a and substantially parallel second planar end surface 132 b withfastener apertures 134 therein. Head spacer 124 may be constructed fromany suitable material including metal or plastic. Cylindrical portions136 define suction passageways 138 therethrough, and are connected tobody portion 130 by bridge portions 140. The remainder of thesubstantially open interior of body portion 130 partially definesdischarge chamber 128, and cooperates with cylinder head 66 to form anenlarged discharge chamber 128. Head spacer 124 thereby cooperates withcylinder head 66 to effectively increase the volume of discharge chamber128 of cylinder head assembly 122, in order to prevent the buildup ofdischarge pressure gas within discharge chamber 128. Discharge chamber128 therefore may accommodate a greater volume of discharge gas,allowing substantially all of the discharge gas to be exhausted fromcylinder bores 40 during the operation of compressor 104, improving theefficiency of compressor 104. When assembling cylinder head assembly122, first and second end surfaces 132 a, 132 b of head spacer 124 aresealed with the adjacent surfaces of cylinder head 66 and valve assembly78, respectively, by gaskets (not shown). Compressor assembly 20 of thefirst embodiment is not provided with head spacer 124 due to a lack ofclearance within housing 22, however, if space were available, headspacer 124 would improve the efficiency of compressor 20 in the samemanner as described above. As noted above, the discharge chamber withinthe head generally acts as an expansion chamber muffler, and enlargementof this chamber generally improves its effectiveness as such.

The second embodiment of head spacer 124′, shown in FIGS. 18C-18E, whichis provided with an alternative sealing method between surfaces 132 a′and 132 b′ of spacer 124′ and valve plate assembly 78 and cylinder head66. Spacer 124′ may be formed of an injection-molded plastic. Theplastic material has a coefficient of thermal expansion which issubstantially similar to the metal components of cylinder assembly 122,including cylinder block 41 and cylinder head 66, such that head spacer124′ may shrink and/or expand at substantially the same rate as cylinderblock 41 and cylinder head 66. The plastic material of which head spacer124′ is formed provides insulation against thermal conduction betweendischarge chamber 128 and suction chamber 126. One suitable plastic forhead spacer 124′ is PLENCO® a phenolic molding compound, Product No.6553, available from Great Lakes Plastics, 7941 Salem Rd., Salem, Mich.,which, after curing, has a coefficient of linear expansion of 12×10⁻⁶mm/mm/° C. (25° C. to 190° C.). (PLENCO® is a registered trademark ofPlastics Engineering Co., 3518 Lakeshore Rd., Sheboygan, Wis.)

Referring to FIGS. 18D and 18E, the alternative method of accomplishingthe above described sealing engagement of head spacer 124′ includesproviding a series or plurality of concentric, continuous ridges 142 onsubstantially parallel planar end surfaces 132 a′, 132 b′ of head spacer124′ disposed around the periphery of body portion 130′ havingcorresponding and alternating ridge tips 144 and valleys 146. As may beseen in FIGS. 18C, ridge tips 144 and valleys 146 are continuous, andcircumferentially extend around the periphery of first and second endsurfaces 132 a′, 132 b′ of head spacer 124′. Referring again to FIG.18D, ridges 142 are shown in an undeformed state, where tips 144 extenda first distance D₁ from each of first and second planar end surfaces132 a′, 132 b′, and valleys 146 extend a second distance D₂ from each ofplanar first and second end surfaces 132 a′, 132 b′ opposite tips 144.As shown in FIG. 18D, first distance D₁ is approximately twice thelength of second distance D₂, but may vary substantially. First andsecond end surfaces 132 a′, 132 b′ lie in planes perpendicular to a lineL₁-L₁, which defines a central axis of head spacer 124′. When headspacer 124′ is placed between valve plate 78 and cylinder head 66 duringassembly of cylinder head assembly 122, and a compressive load isexerted upon cylinder head assembly 122, for example, by torquing downfasteners such as bolts (not shown) to tighten cylinder head 66, ridgetips 144 plastically deform to a deformed state as shown in FIG. 18E.

In the deformed state shown in FIG. 18E, ridge tips 144 are deformed bythe planar interfacing surfaces 148, 150 of cylinder head 66 and valveplate 78, respectively, into a generally mushroom shape in whichportions of ridge tips 144 extend into adjacent valleys 146, andportions of adjacent ridge tips 144 may contact one another to formsealing surface 152 between head spacer 144 and cylinder head 66, aswell as between head spacer 124′ and valve plate 78. Sealing surfaces152, created by the deformation of ridge tips 124′, define labyrinthseals 154. Labyrinth seals 154 are tortuous arrangements of deformedridge tips 144 which seal discharge gas within discharge chamber 128 atthe interface of head spacer 124′ and cylinder head 66, as well as atthe interface of head spacer 124′ and valve plate 78. Labyrinth seals154 sufficiently seal head spacer 124′ between cylinder head 66 andvalve plate 78, obviating the need for additional seals. It may be seenfrom FIG. 18E that the interfacing surfaces of cylinder head 66 andvalve plate 78 respectively lie in first and second planes which arerespectively substantially coincident with the third and fourth planesdefined by first and second end surfaces 132 a′, 132 b′ of head spacer124′, respectively, when the fasteners are tightened to torque cylinderhead 66 down onto head spacer 124′, valve plate 78, and cylinder block41, and causing ridge tips 144 to deform to form labyrinth seals 154.

Generally, during the assembly of compressor 104 and cylinder headassembly 122, cylinder head 66, valve plate 78, and head spacer 124 arepositioned respectively adjacent one another, in a stacked arrangementon cylinder block 41, such that cylinder bores 40 are covered, andfastener apertures 134 in head spacer 124 and the foregoing componentsare aligned. Fasteners are then inserted through apertures 134 incylinder head assembly 122 to engage cylinder block 41 and exert acompressive load on cylinder head assembly 122. This tightens cylinderhead assembly 122 down onto cylinder block 41, which, seals adjacentsurfaces 132 a, 132 b and cylinder head 66 and valve plate 78,respectively. In the case of the alternative sealing method, ridge tips144 of head spacer 124′ are compressed from the undeformed state shownin FIG. 18D to the deformed state shown in FIG. 18E, providing sealingsurfaces 152 and labyrinth seals 154 between head spacer 124′ andcylinder head 66, and between head spacer 124′ and valve plate 78.

The flow of gas through compressor assembly 104 is similar to that ofcompressor assembly 20. The suction pressure gas flows into suctionchamber 126 defined in cylinder head 66 and head spacer 124. Fromchamber 126, the suction pressure gas passes through suction ports 80(FIG. 8A) of valve plate 78 into cylinder bores 40 where the refrigerantis compressed to a substantially higher discharge pressure. Thecompressed fluid flows through discharge ports 81 of valve plate 78 intodischarge chamber 128 also defined by cylinder head 66 and head spacer124. The discharge pressure gas in chamber 128 exits cylinder headassembly 122 through discharge outlet 86 illustrated in FIG. 10 andenters first muffler 156 (FIGS. 10, 11, 12 and 20).

Referring now to FIG. 20, it can be seen that first muffler 156comprises tube 160 having a diameter of approximately ⅝ inch, which maybe a part of discharge tube 88. Tube 160 extends through generallycylindrical shell 162 having first and second ends 164 and 166. Shellends 164 and 166 are sealed to the exterior surface of tube 160 andwithin shell 162, tube 160 is provided with a plurality of holearrangements 168. Each arrangement of holes 168 comprises three pairs ofholes 170, the holes in each arrangement may be cross drilled so thatthe holes are equally radially distributed about the circumference oftube 160. In this embodiment, each hole 170 is formed in the shape of anellipse having an area of 0.0345 square inches. Notably, eacharrangement of holes 168 are substantially equally spaced along thelongitudinal axis of tube 160. It is understood that holes 170 may be ofany shape and size that adequately attenuate noise in the dischargepressure refrigerant.

As with compressor 20, referring now to FIG. 21, there is showndischarge tube 96 which may be part of discharge tube 88 both of whichbeing approximately ½ inch in diameter and constructed from steel.Located in discharge tube 96 is second muffler or resonator 158 as shownin greater detail in FIG. 22. Like the first muffler 156, secondresonator 158 comprises part of a tube which extends through a shell,the tube within the shell having a plurality of spaced holearrangements. As shown in FIG. 22, tube 171 extends through shell 172which has first and second ends 174 and 176. Ends 174 and 176 of thegenerally cylindrical shell 172 are sealed to the exterior surface oftube 171. A plurality of hole arrangements 178 are axially spaced alongtube 171 within shell 172, each arrangement of holes 178 comprising aplurality of holes 180. As described above, holes 180 may be crossdrilled or punched through tube 171, thereby equally radiallydistributing the holes about the circumference of the tube. Holes 180are of similar size and shape to holes 170 of first muffler 156. Secondmuffler 158 is spaced from first muffler 156 along discharge tube 96 aspecific distance to better attenuate noises in the primary pumpingpulse in the discharge pressure refrigerant as will be describedhereinbelow.

It is to be noted that although first and second mufflers 156 and 158depicted are of the Helmholtz type, it is to be understood that thepresent invention may be practiced using first and second mufflers whichare merely expansion chambers. Such mufflers would not have a tubeextending longitudinally through the muffler, but rather would have atube which enters into the expansion chamber, which may be defined byshells 162 and 172, and a tube which exits from the shell, the interiorof the mufflers being open and hollow.

Compression devices such as hermetic compressors 20 and 104 (FIGS. 1, 2,and 10-12) are driven at a particular frequency which correlatesdirectly with the speed at which driving motor 28 disposed withincompressor shell or housing 22 rotates. As described above, motor 28,which is well known in the art, has rotor 34 which iselectromagnetically induced into rotation by current directed throughwindings 33 in stator 32. Shaft 36 extending longitudinally throughrotor 34 drives compression mechanism 30. Thus, the frequency of thepressure pulses will be directly correlated to the speed of motor 28.The speed of motor 28 in compressors 20 and 104 is approximately 3450 to3500 rpm which directly correlated to the frequency of the alternatingcurrent which powers motor 28. Thus, the frequency of the pulse which isassociated with the frequency of the alternating current which powersmotor 28, can be predicted with accuracy because the cycle of theelectrical power is a known quantity. For example, in the United States,electrical power of the alternating current type is normally provided ata 60 hertz cycle.

The cyclical pulsations in the refrigerant which result from itscompression within compression mechanism 30 and which is directly andmost elementally correlated to frequency of the electrical power whichdrives motor 28, may be referred to as the primary pumping frequencywithin the primary pumping pulse. The primary pumping frequency willalso be affected by the number of compression chambers which arecompressing the fluid directed through discharge tube 88. For example, areciprocating piston type compressor may have a single cylinder andpiston. Thus, the primary pumping frequency will be a factor of onetimes the frequency at which electrical power is provided to the motor.Similarly, as is the case with compressors 20 and 104, a reciprocatingcompressor which has two cylinders 40 and pistons 38 driven off commonshaft 36 will have a primary pumping frequency which is twice that ofthe single piston type compressor. Accordingly, a three piston typecompressor will have a pumping frequency which is three times that ofthe single piston type compressor, and so on.

The primary pumping frequency wave form in the primary pumping pulse inthe discharge pressure refrigerant has both a standing or nonmovingcomponent as well as a traveling component, each of which havingdifferent amplitudes to produce different sounds or noises. Theamplitude of the standing wave is much greater than the traveling waveand has fixed peaks and valleys as depicted in FIG. 23B. The travelingwave (not shown) has a much smaller amplitude that produces much lessnoise during compressor operation than the standing wave. The amplitudeof the traveling wave is reduced as the wave moves along a muffler orresonator, no specific placement of the muffler is required because thepoints of amplitude maximum absolute value (i.e., the points of lowestminimum or highest maximum amplitude) of the primary pumping frequencyare not fixed. However, in order to effectively reduce the amplitude ofthe frequency of the standing wave, the muffler must be placed at thefixed points of amplitude maximum absolute value (i.e, the points oflowest minimum or highest maximum amplitude) of the primary pumpingfrequency wave form.

A single Helmholtz muffler is capable of reducing the amplitude of veryspecific frequencies, however, only in a narrow band width. Expansionmufflers are capable of reducing the amplitude of frequencies in a wideband width, however, the amplitudes attenuated are much lower than aHelmholtz resonator. In order to effectively reduce the noise producedduring compressor operation by the primary pumping pulse, a singlemuffler and the compressor discharge chamber, or a pair of mufflers, arespaced along the discharge tube, at specifically calculated points inthe primary pumping frequency wave form as is discussed below.

In accordance with the present invention the first and second mufflersof both compressors 20 and 104 are placed in series along the dischargetube assembly at a specific distance from one another, that distancecorresponding to that distance between the expected minimum and maximumamplitudes of the primary pumping frequency wave form in therefrigerant. In compressors 20 and 104, a problematic or noisy frequencyis produced by a discharge pulse within the primary pumping pulse havinga frequency of approximately 1400 hertz created by operation ofdischarge valve 84 of valve assembly 74. Accordingly, discharge chamber76 in cylinder head 66 and mufflers 100, 156 and 158 are tuned andaxially spaced along discharge tube 88 to reduce the amplitude of thedischarge pulse at a frequency of 1400 hertz. It is understood that themufflers are tuned for the use of refrigerant R22, if an alternativerefrigerant were used in compressors 20 and 104, the mufflers would haveto be retuned.

The first muffler, which is essentially discharge chamber 76 in cylinderhead 66 of compressor 20, and first muffler 156 of compressor 104, whichmay be positioned at any point downstream of head 66, establish aninitial point from which wavelength λ is measured. With reference now toFIG. 23B, wavelength λ is represented by a sine wave which begins atpoint A and ends at point B. Although FIG. 23B shows that point Acoincides with a node or a point of minimum amplitude of the wave, it isto be understood that this placement of the first muffler need not be atsuch a node. In any case, the amplitude of the pressure wave exiting thefirst muffler will be reduced, at that frequency, relative to itsamplitude prior to entering the first muffler. Thus wave form 182extends for one complete wavelength λ between points A and B. Asdepicted in FIG. 23B, where wave form 182 has a node coinciding withpoint A, one half of wavelength λ also occurs at a node, as does thepoint of wave form 182 which coincides with point B. At one quarter andthree quarters the length of wavelength λ from point A, it can be seenthat wave form 182 has maximum amplitudes 184 and 186. Those of ordinaryskill in the art will recognize that at any other odd multiple of onequarter λ, wave form 182 will also be at a point of maximum absoluteamplitude value. As shown in FIG. 23B, distance D is that distance frompoint A to the point of maximum amplitude 184 at one quarter λ anddistance D′ is the distance between point A at maximum amplitude 186 atthree quarter k. These distances D and D′ correspond to the spacingbetween the first and second muffler as illustrated in FIGS. 23A. Themufflers in FIG. 23A are represented as mufflers 156 and 158 ofcompressor 104, however, it is understood that mufflers 76 and 100 ofcompressor 20 could be represented in place of mufflers 156 and 158,respectively. Wave form 182 demonstrates a frequency and generalcharacter of a pressure wave, the relationship between the wave formbeing that frequency of the primary pumping frequency. Thus thestructure of the present invention can be established with help of thefollowing equation:c/f=λwhere c equals the speed of sound in the compressed refrigerant; fequals the primary pumping frequency; and λ is the wavelength.

The operating speed of compressors 20 and 104 running on a 60 hertzelectrical input is 58 hertz. Compressors 20 and 104 being two cylindertype piston compressors, the primary pumping frequency is 2 times 58hertz which approximately equals 116 hertz. This is incorporated intothe above equation. The speed of sound in refrigerant is 7200 inches persecond, however, this may vary with temperature and pressure.

The resulting λ is 62 inches. The point of maximum amplitude 184 at onequarter λ is thus 15½ inches. Thus, in order to further attenuate theamplitude of the pumping pulse in the discharge fluid, second muffler100 or 158 should be located at a distance D of 15½ inches from firstmuffler 76 or 156, respectively. Alternatively, second muffler 100 or158 can be located at distance D′ from first muffler 76 or 156, thisdistance corresponding to three quarters of the length λ or 46½ inches.Thus, by means of the present invention, the second muffler, by beingplaced at a particular distance corresponding to points of maximumamplitude of the pressure pulses in the primary pumping frequency, fromthe first muffler, the noise associated with the primary pumpingfrequency can be effectively and further attenuated vis-a-vis previoussystems having but a single discharge muffler. The two mufflers of eachcompressor do not necessarily have to be precisely placed at 15½ inchesfrom each other and may be placed a distance of approximately 12-20inches apart before reaching a higher discharge pulse near a node.

With reference to mufflers 156 and 158 of the Helmholtz type, as shownin FIG. 23A, distances D and D′ shall be most effectively extended fromthe furthest downstream arrangement of holes 170E in first muffler 156and furthest upstream arrangements of holes 180A in second muffler 158.By so arranging the first and second Helmholtz type mufflers 156 and158, the greatest attenuation of the primary pumping pulse can beachieved by the first muffler, the second muffler having the greatestopportunity then to further attenuate the pumping pulse which reachesit.

Referring again to FIG. 23B as discussed above, wave form 182 representsa sine wave, which may be representative of the pressure pulse betweenthe two mufflers, demonstrating the wavelength and points of maximumamplitude 184 and 186 along wavelength λ. The diminishing wave form isfurther shown in FIGS. 23B has a first amplitude A1 before enteringfirst mufflers 76 and 156. After passing through the first mufflers, theamplitude of waveform 182 at point 184 is reduced at 188 to having anamplitude of A2 (FIG. 23B). With second mufflers 100 and 158 located atdistance D from first mufflers 76 and 156, respectively, it can be seenin FIG. 23C that wave form 182 will enter second mufflers 100 and 158having an amplitude of A2 and will be reduced as at 190 to having anamplitude of A3 upon exiting the second mufflers. Similarly, with secondmufflers 100 and 158 located at a distance D′ corresponding to point ofmaximum amplitude 186, at three quarter λ, it can be seen that theamplitude A2 of wave form 182 will be reduced as the refrigerant passesthrough second mufflers 100 and 158, to a modified wave form shown at190 having a reduced amplitude A3 (FIG. 23B).

Although compressors 20 and 104 depict that first muffler 76, 156 andsecond mufflers 100, 158 are packaged within housing 22, it is to beunderstood that the separation of the first and second mufflers may beachieved in a discharge line external to housing 22. The placement ofthe first and second mufflers along discharge tube 96 within housing 22improves the packaging characteristics of compressors 20 and 104, but isnot a necessary aspect of the present invention.

During the operation of compressor assemblies 20 and 104, thecylindrical shape of housing 22 has several natural resonant frequenciesthat produce loud, pure tones which are undesirable. In order to reduceor eliminate these frequencies, resilient mounts 24 illustrated in FIGS.5 and 16 are welded to housing 22 so as to span a node and an anti-nodeof the wave form. Mounts 24 are secured at 196 to crankcase 42 and at198 to the inner surface of housing 22 by means such as weldment. Thenatural frequencies associated with housing 22 may have any number ofnodes. The most problematic or noticeable frequency 193 is one in whichthere are six naturally occurring nodes 192 and anti-nodes 194circumferentially spaced around housing 22 at equal distances (FIG. 24).

To reduce the amount of noise produce by this natural frequency, thenodes and anti-nodes must be forced to an alternative position byspecifically securing mounts 24 to housing 22 at points which areunequally distributed about the circumference of housing 22 and which donot coincide with naturally occurring nodes. The forced frequency 193′produced by mounts 24 is illustrated in FIG. 25 and is represented bydashed lines. It is critical that mounts 24 are unequally distributedabout the circumference of housing 22 because if they were equallydistributed, forced nodes 192′ and anti-nodes 194′ would fall on thoseof natural frequencies and thus the amplitude of the natural frequencywould not be attenuated.

Referring to FIG. 25, one of ends 198 of each mount 24 is welded to theinside surface of housing 22 at positions offset from naturallyoccurring nodes 192. The weld forces nodes 192′, dampening thevibrations in housing 22 created by the natural frequency. The weld atopposite end 198 of mount 24 is then located so as to force anti-node194′ or points of maximum amplitude between two nodes. Forced anti-nodes194′ are then free to vibrate and cause tones which produce noise. Thesetones, however, are at a much lower amplitude which do not produce thesame objectionable noise of the natural resonant frequencies.

While this invention has been described as having exemplary designs, thepresent invention may be further modified within the spirit and scope ofthis disclosure. Therefore, this application is intended to cover anyvariations, uses, or adaptations of the invention using its generalprinciples. For example, aspects of the present invention may be appliedto compressors other than reciprocating piston compressors. Further,this application is intended to cover such departures from the presentdisclosure as come within known or customary practice in the art towhich this invention pertains.

1-4. (canceled)
 5. A method for reducing the amplitude of pressurepulses having a particular wavelength in a fluid, comprising: flowingthe pressure pulse-containing fluid through a conduit; attenuating thepressure pulse amplitude at a first location along the conduit; andfurther attenuating the pressure pulse amplitude at a second locationalong the conduit distanced from the first location a distance which issubstantially equal to an odd multiple of one quarter of the wavelength.6. The method of claim 5, wherein said steps of attenuating and furtherattenuating respectively comprise flowing the fluid through a firstmuffler and a second muffler, the first and second mufflers respectivelylocated at the first and second locations.
 7. The method of claim 6,wherein said step of flowing the fluid through a first muffler comprisesflowing the fluid through a head discharge chamber of a reciprocatingpiston compressor.
 8. The method of claim 6, wherein at least one ofsaid first and second mufflers is a Helmholtz muffler.
 9. The method ofclaim 5, wherein said step of flowing the fluid through a conduitcomprises flowing the fluid through the head discharge chamber of areciprocating piston compressor, and said step of attenuating thepressure pulse amplitude at a first location comprises flowing the fluidthrough the head discharge chamber. 10-28. (canceled)